It is desired that fluid pressure regulators coupled to compressed gas cylinders or other containers containing pressurized gas at 2000 psi or greater produce a constant outlet pressure as the gas in the cylinder is used up, and the cylinder pressure decreases. For example, constant outlet pressures for pressurized nitrogen, nitrogen-arsine/phosphorous gas mixtures as dopants, argon, hydrogen and air, are desired in the manufacture and treatment of semiconductors. However, in typical known single-stage fluid pressure regulators, the outlet pressure increases by 9 psi as the cylinder pressure decreases from 2000 psi down to 200 psi. A variation in outlet pressure of 9 psi is relatively substantial in relation to the typical outlet pressure of the regulator, which may be in the range of 25 to 35 psi. This is a problem where process requirements, as in semiconductor manufacture, call for constant or substantially constant gas pressure.
A prior art, single-stage fluid pressure regulator 1 is schematically illustrated in FIG. 1 of the drawings. The pressure regulator 1 is shown connected at its inlet to a source of gas at high pressure, particularly a gas cylinder 2, which initially, when filled, contains a gas at a pressure of 2000 psi or higher. The function of the regulator is to receive pressured gas from the cylinder at the regulator inlet, which is connected to the cylinder, and to deliver the gas at a selected lower pressure from the outlet of the regulator, while minimizing the change in outlet pressure in the presence of changes in outlet flow and variations in supply pressure.
The regulator 1 comprises a diaphragm 3 and a valve 4. One end of the valve, the upper end in FIG. 1, contacts the diaphragm; and the other, outwardly tapered end of the valve cooperates with valve seat 5 for adjustably throttling a fluid passage 6 extending through the regulator. Particularly, the fluid passage 6 extends through central aperture 7 of the valve seat 5. The upper surface of the annular valve seat 5 is supported about the aperture 7 by regulator member 8. The lower, conical portion 9 of the valve 4 cooperates with the lower edge 10 of the seat about central aperture 7 for adjustably throttling the fluid passage through the regulator, thereby controlling the outlet pressure from the regulator. A spring 11 applies an adjustable force to the diaphragm 3.
The pressure regulator 1 operates on a force balance principle. The diaphragm 3 has an effective area A. The upper side of the diaphragm 3 is exposed to atmospheric pressure and has force F.sub.S applied thereto by the spring 11. The force F.sub.S is balanced by the force F.sub.Po created by the outlet pressure P.sub.o. that is the fluid pressure downstream of the throttled aperture 7 in valve seat 5, which acts on the diaphragm and a force F.sub.Ps, created by the supply pressure P.sub.s, that is the fluid pressure from the gas cylinder 2 in the regulator 1 upstream of the throttled aperture 7 of the valve seat 5, acting against the valve 4 on an area a determined by the size of the aperture 7, geometry of valve, and size of opening.
Any imbalance will result in a deflection of the diaphragm 3 to vary the size of the flow passage between the valve seat and the lower, conical portion 9 of the valve, that is to vary the amount of throttling of fluid passage 6 through the regulator 1, and thereby create a new value of P.sub.o in order to restore the balance. If there is an increase in outlet flow, P.sub.o starts to decrease. This increases the size of the adjustable passage between valve and valve seat 5 to supply a greater flow. Thus, a small decrease in outlet pressure provides the larger flow demand. If there is a decrease in the supply pressure (pressure decay at the supply cylinder as a result of usage), the force imbalance tends to increase the size of opening between the valve seat 5 and valve 4 which increases P.sub.o and restores the balance. Thus, decrease in supply pressure increases the outlet pressure, for the same flow demand.
An equation defining the operation may be expressed as follows: EQU F.sub.s =F.sub.Po +F.sub.Ps
or EQU F.sub.s =P.sub.O A+(P.sub.s -P.sub.o)a
or solving for the outlet pressure P.sub.o, ##EQU1##
From the above analysis, it is seen that the effect of variations in the supply pressure on the outlet pressure P.sub.O are a function of the expressions: ##EQU2##
In a typical prior art regulator of the type illustrated in FIG. 1, ##EQU3##
resulting in a supply pressure of +0.5 psi per 100 psi decrease in the supply pressure P.sub.s.
The motion of the valve .DELTA.d to change the flow is defined by the equation: ##EQU4##
where k is the spring constant, 1b/inch, of the system (K.sub.s for the spring+K.sub.D for the diaphragm). As the flow is increased, an increase, .DELTA.d, in the valve opening is required to provide the additional flow. This is accompanied by a decrease in outlet pressure P.sub.O.
FIG. 2 illustrates flow curves for the conventional regulator 1 which show outlet pressure P.sub.O as a function of the flow rate for different values of supply pressure P.sub.s. FIG. 3 shows the effect of supply pressure change for regulator 1 with a typical supply pressure effect of 0.5 psi per 100 psi. The slope of the basic supply pressure effect, shown as a dotted line, is modified by the corresponding flow curve as an increase in the supply pressure decreases the valve motion drop for a given flow, but only partially recovers the decrease created by the supply pressure effect. Thus, this conventional, single-stage regulator 1 is undesirable for fluid pressure regulation where a constant or substantially constant outlet pressure for a given flow rate is necessary as the pressure in the gas cylinder decreases from 2000 psi to, for example, 200 psi. From FIG. 3, for example, with a flow rate of 50 liters per minute, it is seen that at 2000 psi, the outlet pressure is between 17 and 18 psi, whereas at 500 psi, the outlet pressure has increased to between 23 and 24 psi, an increase of approximately 8 psi or almost 50 percent.
One prior art solution to this problem is the two-stage fluid pressure regulator. The first stage of the regulator reduces the high pressure from the supply to an intermediate pressure, for example, 400 psi, and a second stage of the regulator further reduces the pressure to a substantially constant outlet pressure. However, two-stage regulators are disadvantageous in that they are relatively costly and heavy.
Another prior art solution to the problem of supply pressure effect in regulators is depicted in FIG. 4, wherein regulator 12 is provided with a balanced poppet 13. A piston 14 opposite the poppet 13 and connected thereto has an area approximately the same as the area of the poppet so that the inlet pressure force acting on the poppet has little effect. However, these regulators can be disadvantageous in that they are relatively large and heavy. The seal or packing 15 about the piston 14 can also be a source of particulate contamination to the pressurized gas. The necessary loose fit of the piston 14 within the regulator for proper functioning is can also result in gas leakage past the seal 15 and the piston 14 in the lock-up state (no-flow condition) of the regulator leading to leakage pressure build up at the outlet of the regulator.
There is a need for an improved fluid pressure regulator which can overcome the problem of the supply pressure effect in an efficient and cost-effective manner, without the aforementioned drawbacks of the conventional solution thereto.